The present invention relates generally to single or multistage axial flow compressors and more particularly to a device for extending the working range thereof.
Generally, the working range of single or multistage axial flow compressors during throttling with decreasing mass throughput is limited due to the so-called pumping phenomenon which occurs. In an axial flow compressor, the flow will tend to become turbulent and completely break away. Furthermore, it is known that axial flow compressors, prior to reaching the pumping limit, operate in a range in which flow breaks away partly on the blades. Also, so-called rotating stalls are then formed. Due to the occurrence of the rotating stalls, corresponding axial compressor blades experience vibrations which under certain circumstances can lead to failure of the blades. Depending upon the throttling stage and the speed ratio, there may occur up to eight rotating stalls which under certain circumstances could be capable of causing blade ruptures. Examinations of blade vibrations caused by rotating stall have also shown that alternating bending stresses are developed which amount to between about 2.5-3 times the bending caused by gas forces under static conditions. The blade materials must thus meet much higher requirements when rotating stall is present.
In order to avoid such effects of rotating stall, it has been suggested to provide binding or damping wires in the jeopardized rings of the blade. Such wires are arranged to connect with each other the individual blades of a blade ring in the circumferential direction. Their function is to damp as far as possible blade vibrations caused by rotating stall, or to eliminate such vibration completely in order to avoid vibration failure of the blades.
A disadvantage of utilizing such damping wires arises as a result of the fact that the wires introduce additional flow resistances into the compressor. For example, for a multistage flow compressor, the efficiency may be reduced by as much as 5% for a damping wire.
Another approach which has been known for some time involves the use of suitable blow-off or reblowing controls. It has been known that the additional stresses on the blades of axial flow compressors, as they appear in pumping and in rotating stall, can be avoided by adapting the working range of the axial flow compressor by means of such blow-off or reblowing controls to the changing reduction. However, utilization of these methods has been found to be relatively uneconomical since only a part of the throughput is delivered to the load while a a remaining part is reblown or blown off.
In addition to fixed blow-off or reblowing controls, which constantly limit the characteristic fluid, there is known a safety control device which acts only when rotating stall appears. This device first determines the appearance of rotating stall by means of a Pitot tube arranged behind a runner in the circumferential direction. When rotating stall appears, a blow-off valve arranged in the pressure line is opened over a control device. The valve closes again after the compressor has entered the stable working range. This safety device differs from presently known pumping prevention devices in that the blow-off valve opens automatically in the case of rotating stall.
However, devices of this type exert no influence on the causes of rotating stall. Furthermore, other disadvantages arise due to the fact that the device is relatively elaborate.
In addition to the foregoing, it has also been suggested that shroud bands be utilized in order to avoid blade vibrations in axial flow compressors. Although the use of shroud bands or guide wheels is known, shroud bands cannot be used for the blade wheels of axial flow compressors operating at contemporary circumferential speeds because of the appearance of high centrifugal forces. With binding or damping wires, or with shroud bands, it is possible to avoid blade vibrations caused by rotating stall without influencing the formation of the rotating stall. It has also been suggested to shield, at a relatively early stage, the outer return flow region which appears during pumping by arranging a concentric non-airfoil ring (i.e., a ring not having an airfoil cross section relatively close to the compressor casing wall in front of a blade wheel in order to influence the location of the pumping limit. A disadvantage of this approach involves the fact that the approaching flow direction to the following blade wheel is not changed so that the appearance of rotating stall cannot be avoided. Initially, it must be taken into consideration that flow breaks away at the non-airfoil ring and thus rather enhances the formation of rotating stalls. Indeed, it is likely that flow will stall on the non-airfoil ring and there will rather be an increase in the formation of rotating stall. Furthermore, the flow stalled on the non-airfoil ring causes considerable losses so that machine efficiency is substantially reduced for the axial flow compressor.
On the other hand, added loading upon the blading of axial flow compressors, as might occur in pumping and in rotating stall, is avoided since the working range of the compressor is limited by means of a suitable blow-off or by deflection regulators. However, it becomes apparent that both methods are relatively uneconomical, since only a part of the delivered throughput is applied to the load while the remaining portion is deflected or blown off.
There are also known proposals to arrange so-called profiled hub rings in the vicinity of the hub to avoid hub detachment in axial flow machines. However, since these rings only influence the formation of hub detachments, they cannot avoid the formation of rotating stall. Also, the hub ring cannot achieve change in the overflow direction of the succeeding wheel, inasmuch as the ring surfaces are the same on the inlet and outlet side of the hub ring. Thus, a converging annular wind tunnel is not formed in the range of the casing wall.
With the aforementioned considerations, it is only possible to maintain the effects of rotating stall within certain limits. The causes of rotating stall, which arise from the fact that the aerodynamic load capacity of the cascades is partially exceeded, cannot be influenced.
The present invention is essentially directed toward avoiding the aforementioned disadvantages and toward extending the working range of axial flow compressors toward lower output volumes with minimum additional flow losses. This is accomplished in the present invention by utilization of a suitable device whereby the occurrence of rotating stall is, as far as possible, prevented.